Fuel Processing Technology 133 (2015) 20–28
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Fuel Processing Technology journal homepage: www.elsevier.com/locate/fuproc
Combustion, performance and emission characteristics of fusel oil in a spark ignition engine Hamit Solmaz ⁎ Automotive Engineering Department, Faculty of Technology, Gazi University, Ankara, Turkey
a r t i c l e
i n f o
Article history: Received 31 October 2014 Received in revised form 8 January 2015 Accepted 9 January 2015 Available online 20 January 2015 Keywords: Fusel oil Spark ignition engine Exhaust emissions Combustion Alternative fuel Alcohol
a b s t r a c t Alcohol based fuels attract the attention of alternative fuel researchers. Many studies have been performed about combustion, performance and emission characteristics of alcohol used in internal combustion engines. Fusel oil is an alcohol based fuel obtained as a by-product during alcohol fermentation. Up to the present there has been no study regarding the combustion characteristics of fusel oil in a spark ignition engine. In this experimental study, performance, emission and combustion characteristics of fusel oil were examined in a spark ignition engine at 2500 rpm and four different engine loads. In-cylinder pressures, heat release rates, ﬂame development and ﬂame propagation durations, crank angles corresponding 50% of total mass fraction burnt, engine torque, brake speciﬁc fuel consumptions, CO, HC and NOx emissions were investigated. The water content and lower heating value of the fusel oil aggravated the combustion. Flame development and ﬂame propagation durations were prolonged. As a result engine performance dropped. In addition, fusel oil usage increased CO and HC emissions up to 21% and 25% respectively. NOx emissions decreased about 31% due to worse combustion performance of fusel oil. © 2015 Elsevier B.V. All rights reserved.
1. Introduction A large part of fuel used in the motored vehicles is fossil fuels. However researchers have studied on the alternative fuels owing to both damage on the environment and depletion of fossil fuels [1–3]. Alternative fuels should be environmental, renewable and easily obtained energy source. In addition, they can be used with minimum modiﬁcations in the internal combustion engines. It is also possible to say that alternative fuels can improve the combustion and engine efﬁciency. As known fossil fuels lead to global warming because of their high carbon dioxide (CO2) release. Besides, carbon monoxide (CO) and hydrocarbon emissions (HC) have carcinogenic and toxic effects that are released when the spark ignition engines operate with inappropriate air/fuel ratios . The reduction of these harmful exhaust emissions depends on the providing the appropriate combustion conditions in a spark ignition engine. Fuels such as ethanol and methanol improve the combustion process and CO emissions are reduced especially [5,6]. Engine efﬁciency is determined by compression ratio in the internal combustion engines. The ability to increase the compression ratio depends on the octane number of fuel used in the spark ignition engines. Nowadays, limiting the octane number at about 100 prevents the increase of compression ratio in the spark ignition engines. For this reason, the efﬁciency of spark ignition engines are lower than compression ignition engines. High octane number fuels or improver ⁎ Tel.: +90 312 2028646; fax: +90 312 2028947. E-mail address: [email protected]
http://dx.doi.org/10.1016/j.fuproc.2015.01.010 0378-3820/© 2015 Elsevier B.V. All rights reserved.
additives which increase the octane number is essential in order to prevent knocking problem occurred at higher compression ratios [7–12]. Iodine, tel tetra-ethyl lead and alcohol-based additives have been used in order to increase the knocking resistance until now. The researches on the alcohol-based additives increased since the usage of tetra-ethyl lead is forbidden due to harmful effects on health and iodine damages the engine parts [13,14]. Many researchers have studied on usage of alcohols directly as an alternative fuel and fuel additives in the spark ignition engines. Lower heating values of alcohols are lower than gasoline. Therefore, fuel consumption usually increases when alcohol is used as an alternative fuel [15,16]. Bata et al.  concluded that the ethanol addition to gasoline reduced the CO and HC emissions. Similar results were obtained in many researches [5,6,18–22]. The reduction in CO and HC emissions was caused by oxygenated characteristic and wide ﬂammability of ethanol. Additionally, similar CO and HC emission reduction trends were seen in use of methanol with gasoline [23–29]. Taljaard et al. conducted a study to investigate the effects of oxygenate in a spark ignition engine. They examined engine performance and exhaust emissions in a single cylinder for stroke engine. It was reported that the CO, NOx and HC emissions reduced signiﬁcantly at stoichiometric air/fuel ratio when the oxygenates were used . Bilgin and Sezer  reported an increase in brake mean effective pressure with 5% methanol addition to gasoline. Hsieh et al.  investigated the effects of 10%, 20% and 30% ethanol–gasoline fuel blends on engine performance in a spark ignition engine. They concluded that the engine torque and fuel consumption slightly increased with ethanol–gasoline blends. They depicted that
H. Solmaz / Fuel Processing Technology 133 (2015) 20–28
the increase in engine torque and fuel consumption could result from improved combustion efﬁciency and lower caloriﬁc value of ethanol respectively. Similar results were obtained by Al-Hasan . He performed the experiments in a four cylinder four stroke spark ignition engine. He investigated the effects of ten different ethanol–gasoline blends on engine performance and exhaust emissions. It was reported that the engine power, brake thermal efﬁciency and volumetric efﬁciency were increased averagely by 8.3%, 9% and 7% respectively. It was reported that the volatility and the latent heat of fuel blend increased while the percentage of ethanol increased in fuel blend. Therefore, charge temperature decreased and volumetric efﬁciency increased [20,32]. Liu et al.  performed a study in a spark ignition engine fuelled with methanol/gasoline fuel blends. They determined start of combustion (SOC) and rapid burning phase (RBP) in case of methanol addition to gasoline. They deﬁned the SOC and RBP from crank angle corresponding 5% and 5%–90% accumulative heat release respectively. They concluded that the SOC was advanced and RBP became shorter with using methanol. A similar result was reported by Hu et al.  and Yanju et al. . They were also concluded that the peak cylinder pressures increased when methanol–gasoline blends were used. Bielaczyc et al.  investigated the effects of the blends of ethanol and gasoline (from 5% ethanol–95% gasoline to 50% etanol–50% gasoline in vol.) on the engine performance and exhaust emissions. The experiments were conducted with an unmodiﬁed European passenger car on a chassis dynamometer over new European driving cycle. The authors presented the results of both regulated and unregulated emissions. It was reported that the lowest HC and CO emissions were observed with E50 due to improved combustion or improved removal in the after treatment system. Also, the lowest NOx emission was observed with E25 in urban driving cycle. Agarwal et al.  investigated the effects of M10 (10% methanol and 90% gasoline) and M20 (20% methanol and 80% gasoline) the blends of methanol and gasoline fuels on the performance, emissions and combustion in a medium duty spark ignition engine. The test results showed that thermal efﬁciency obtained by the blends of methanol– gasoline test fuels were higher than gasoline. They also determined that CO, nitrogen oxides (NOx) and soot emissions decreased when compared to gasoline fuel. It was also found that there was a slight difference in cylinder pressure compared to the gasoline fuel. They noticed that the heat release rate obtained with gasoline started to increase earlier compared to the blends of methanol–gasoline test fuels. Moreover, it was shown that combustion duration decreased with increasing engine load. Maurya and Agarwal  examined the ethanol, methanol and gasoline fuels in a four stroke, port type fuel injection system homogeneous charged compression ignition (HCCI) engine and investigated the performance, emissions and combustion characteristics. They investigated the effects of inlet air temperature and air/fuel ratio on thermal efﬁciency, combustion efﬁciency and emissions. It was shown that ethanol and methanol fuels could be used instead of gasoline in HCCI combustion mode. They realized that the ethanol and methanol fuels could be ignited at lower inlet air temperature compared to gasoline. They obtained higher indicated mean effective pressure (imep) with ethanol and methanol fuels at all constant air/fuel ratio. Siwale et al.  examined the performance and combustion characteristics of gasoline, M53b17 (53% methanol, 17% n-butanol and 30% gasoline in vol.), M20 and M70 test fuels in a spark ignition engine. They pointed out to increase the thermal efﬁciency but decrease the exhaust gas temperature with blends. They saw that the combustion duration decreased with M53b17 due to higher energy content. Furthermore, it was determined that CO emissions increased with M53b17 test fuel compared to M70 test fuel. Fusel oil is obtained as by product in the production of alcohol after the distillation process. Fusel oil has a bad smell and dark brown color. It consists of about 390 g/L isoamyl alcohol, 158 g/L isobutyl alcohol, 28.4 g/L ethyl alcohol, 16.6 g/L methyl alcohol and 11.9 g/L n-propyl
alcohol . It also includes the aldehite, esters and water by about 15% in vol. [33,36–42]. The ﬁrst study on fusel oil was performed by Wetherill in 1853 . Reduction of the harmful effects of fusel oil, reduction of the fusel oil in the alcohol drinks, the lubricating production from the fusel oil and biodiesel production with fusel oil are the researches that have been conducted on the fusel oil up to the present [44–47]. Fusel oil has not been used effectively apart from the compensating the small part of energy demand in the factories. In Turkey 0.4–0.7 L fusel oil is obtained per 100 L alcohol production . According to data given by Turkish Tobacco and Alcohol Market Regulatory Authority, approximately 73,140,000 L ethanol was produced in 2013 in Turkey . It equals about 512,000 L fusel oil production. When it is taken into consideration that the fusel oil has not been used effectively, it is obvious there will be enormous environmental pollution. Using fusel oil even only in agricultural activities will reduce this pollution and cause to decrease cost of agricultural production. There are limited numbers of studies on the usage of fusel oil in the spark ignition engines in the literature. Calam and İçingür  investigated the effects of the blends of fusel oil and gasoline fuels on performance and exhaust emissions. They obtained the maximum engine brake torque with F30 test fuel (30% fusel oil and 70% gasoline test fuel in vol.). Speciﬁc fuel consumption increased by the increase of fusel oil fraction in the test fuels at all engine speed and full load. The highest increase was obtained by F30 test fuel about 7.7%. NOx emissions decreased with the increase of fusel oil fraction in the test fuel but HC and CO emissions increased in the experiments. They determined that the reason of increase of HC and CO was the decrease of the in-cylinder temperature when fusel oil was used. In another study, Calam ve et al.  investigated the effects of the variations of ignition timing on the usage of fusel oil. They saw that engine brake torque and fuel consumption increased by the addition of fusel oil to gasoline in the experiments conducted at 3500 rpm engine speed. In addition, HC and CO emissions increased by about 40% and 10% respectively when fusel oil was used. The usage of fusel oil in the internal combustion engines as an alternative fuel may be advantageous in terms of using a new energy source for internal combustion engines. However, a detailed study on the combustion characteristics of fusel oil is not be found in literature. In this study, the effects of the blends of fusel oil–gasoline test fuels (F0, F50 and F100) on engine performance, exhaust emissions and combustion characteristics were investigated in a single cylinder, four stroke gasoline research engine having port fuel injection system. The experiments were performed at 2500 rpm engine speed, λ = 1 and four different engine loads (25%, 50, 75% and 100%). The effects of test fuels on cylinder pressure, heat release rate, maximum pressure rise rate, combustion durations, engine brake torque, speciﬁc fuel consumption and exhaust emissions were investigated. 2. Experimental apparatus and procedure 2.1. Research engine test bed and test fuels In this study single cylinder, four stroke, port fuel injected gasoline research engine Ricardo Hydra was used. The technical speciﬁcations of the test engine are given in Table 1. Premixed gasoline and fusel oil were provided by port type fuel injector mounted in the intake manifold. Fuel injection pulse width was controlled electronically to keep λ = 1 at constant value. Compression ratio of the test engine can be adjusted between 5:1–13:1. Also, the ignition advance can be switched between 70° before top dead center (BTDC) and 20° after top dead center (ATDC). In this study, experiments were conducted at 9:1 compression ratio and the ignition advance was ﬁxed to 20° BTDC. The test system was also equipped with torque measurement, exhaust gas temperature, ignition timing, injection pulse, coolant and engine oil temperature, air mass ﬂow meter, intake air heater, DC dynamometer and exhaust gas analyzer. Meriam Laminar Flow Element Z50MC2-4F
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Table 1 The technical speciﬁcations of the test engine. Speciﬁcations Model Type Cylinder number Bore × Stroke [mm] Swept volume [cm3] Maximum engine speed [d/d] Maksimum power [kW] Compression ratio Fuel injection system Valve arrengement Valve lift [mm]
Ricardo Hydra Water cooled, natural aspirated 1 80.26 × 88.9 540 5400 15 (at 4500 rpm) 5/1–13/1 PFI Overhead cam, two valve 5.5
and Merriam LFS-1 were used to measure mass ﬂow of the intake air. The intake air temperature was measured with a K-type thermocouple from the entrance of the intake port. The intake air temperature was kept constant at 25 °C by a closed-loop controller. The Farnam Flow Torch 400 electrical heater was used to heat the intake air and ENDAETC9420 PID controller was used to control the air temperature. DC Dynamometer which is McClure brand is able to absorb 30 kW at 6500 rpm engine speed. Dynamometer can be also operated as motor. Engine load can be altered by controlling the electric resistance of dynamometer from control panel. Engine torque was measured by a load cell. Before the experiments, the engine was heated up and coolant (80 °C) and engine oil temperature (70 °C) were kept constant during the experiments in order to eliminate their effects. Exhaust gas was measured by using SUN MGA 1500 exhaust gas analyzer. The technical speciﬁcations of exhaust gas analyzer are given in Table 2. The schematic view of the engine test bed is seen in Fig. 1. The fusel oil used in this study was supplied from Eskişehir sugar reﬁnery which is producing ethyl alcohol with 99.5% purity. Any chemical or physical processes were not conducted on the fusel oil used in this study. The fusel oil was used as it supplied from the reﬁnery. In the experiments three different fuels F0 (100% gasoline by vol.), F50 (50% fusel oil, 50% gasoline by vol.) and F100 (100% fusel oil by vol.) were used in order to determine the effects of fusel oil on performance, combustion characteristics and exhaust emissions. Fuel blends were kept waiting a day after mixed. At the end of twenty four hours there was no any phase precipitation. The properties of the test fuels are given in Table 3. The lower heating value (LHV) of the fusel oil is lower than gasoline like the other alcohol fuels. However, it is seen that the octane number of the fusel oil is higher than gasoline. 2.2. In-cylinder pressure measurement and data processing In-cylinder pressure was measured by Kistler 6121 piezo-electric pressure transducer. Kistler 6121 pressure transducer is able to measure in-cylinder pressure in a range of 0–250 bar and its sensivity is 14.7 pC/bar. Its accuracy is ± 0.5% and working temperature range is 50–350 °C. Cussons P4110 combustion analysis device was used to amplify the analog in-cylinder pressure signals. The analog in-cylinder pressure signals were converted to digital signal with Natural Instruments USB 6259 data acquisition card. In-cylinder pressure signals were recorded with a resolution of 0.36 °CA. Crank angle (CA) and top Table 2 Technical speciﬁcations of exhaust gas analyzer. Products
CO HC NOx CO2 02 λ
0–14% 0–9999 ppm 0–5000 ppm 0–18% 0–25% 0–4
0.001% 1 ppm 1 ppm 0.1% 0.01% 0.001
dead center (TDC) positions were determined with a shaft encoder producing 1000 pulses per revolution. Mean cylinder pressure were calculated by averaging the sampled pressure data of 50 consecutive cycles. Averaged pressure raw data ﬁltered by six point data weighting [50,51]. The heat release rate was calculated according to the ﬁrst law of the thermodynamics by using the equation, dQ k dV 1 dP dQ ht ¼ P þ V þ dθ k−1 dθ k−1 dθ dθ
where dQ is the net heat release. P, V, dθ and k are cylinder pressure, cylinder volume, variation of crank angle and the ratio of the speciﬁc heats respectively. The last term of the Eq. (1) deﬁnes the heat transfer to the cylinder walls. Heat transfer can be calculated by Newton's cooling law, dQ ht 1 hg A T g −T w ¼ 6n dθ
where hg is the convection heat transfer coefﬁcient, A is cylinder wall surface area, Tw temperature of the cylinder wall, n is engine speed and Tg is instant in-cylinder gas temperature. To calculate convective heat transfer coefﬁcient, Woschni was deﬁned an equation based on cylinder bore, in-cylinder pressure, gas temperature and mean gas velocity [52,53]. Hohenberg determined that the heat transfer values using Woschni correlation were low at low engine loads. Hohenberg described a developed equation of Woschni's heat transfer coefﬁcient equation, based on cylinder volume instead of cylinder bore. It is seen that Hohenberg's correlation gives more favorable results for the calculation of heat transfer [42,54–58]. In this study Hohenberg's correlation, hg ¼ 130 V
was used to calculate convective heat transfer coefﬁcient where is cylinder volume and w is mean gas velocity. The following equation can be used for prediction of the in-cylinder temperature. T giþ1 ¼ T gi
Vi V iþ1
where nc is the polytrophic index. Indicated mean effective pressure was calculated by, imep ¼
where Wc is net work and Vd is stroke volume of the engine. In the combustion analysis of the test fuels, ﬂame development and propagation durations, combustion duration and mass fraction burnt were examined. A sample cumulative heat release curve is seen in Fig. 2. CA10, CA50 and CA90 deﬁne the crank angles ATDC corresponding mass fraction burnt rates 10%, 50% and 90% respectively. Flame development duration (CA0–10) was determined by crank angle between spark ignition and CA10. Flame propagation duration (CA10– 90) was calculated by taking the difference CA90 and CA10 points. Knocking tendency during the combustion process was determined from maximum pressure rise rate (MPRR). In Fig. 3 in-cylinder pressure and pressure rise rate (PRR) curves are seen. PRR can be calculated from the ﬁrst derivative of in-cylinder pressure. The maximum point on the PRR curve gives the MPRR. 3. Results and discussions 3.1. Combustion characteristics In Figs. 4–7, the variations of in-cylinder pressures, heat release rates and imeps of F0, F50 and F100 fuels at 2500 rpm engine speed and four
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Fig. 1. The schematic view of the engine test bed. 1. Research engine 2. DC dynamometer 3. Port fuel injector 4. ECU 5. Precision scale 6. Inlet air heater 7. Laminer air ﬂow meter 8. Incylinder pressure sensor 9. Combustion analyzer 10. Encoder 11. Computer 12. Data acquisition card 13. Exhaust gas analyzer 14. UEGO sensor 15. Dynamometer control panel 16. Battery 17. Lambda indicator.
different engine loads (25%, 50%, 75% and 100%) are seen. Experiments were conducted at λ = 1 and 20° BTDC ignition advance condition for all test fuels. Minimum imep, in-cylinder pressures and heat release values were obtained at 25% load due to a decrease in the amount of cylinder charge. It is seen that the in-cylinder pressure and heat release rates increase as engine load increases. The maximum imep value was obtained as 8.05 bar at 100% engine load with F0 fuel. At 25% engine load, differences in in-cylinder pressure and heat release rates of test fuels were very small. However, as engine load in-cylinder pressures increased, heat release rates and imeps of F50 and F100 fuels decreased dramatically compared to F0. This situation indicates that the fusel oil worsens combustion at the same λ and ignition advance. In general, alcohol-based fuels have a positive effect on combustion in internal combustion engines despite their lower LHVs. Alcohol-based fuels are oxygenated fuels that enable more complete combustion. In addition, they provide a higher volumetric efﬁciency in natural aspirated engines due to their higher latent evaporation heat. Fusel oil is an alcohol-based fuel, as speciﬁed above. The worse combustion characteristic of fusel oil is due to its high water content (10–15%). Therefore, when using fusel oil, the maximum in-cylinder pressures and heat release rates decreased. A signiﬁcant reduction was seen in in-cylinder pressure and heat release rates of F100 compared to F0 at 100% engine load. This shows the effect of the water content of fusel oil. However, because of the higher octane number of the fusel oil, the knocking resistance of the engine enhanced. Especially at full load condition F0 has a knocking tendency as seen in Fig. 7. When F50 fuel was used, the knocking tendency largely decreased and completely disappeared by using F100
fuel. This advantageous property of fusel oil may provide better combustion characteristic by determining optimum ignition advance. Fig. 8 shows maximum pressure rise rates of F0, F50 and F100 at four different engine load and 2500 rpm engine speed. MPRR values were determined from the ﬁrst derivative of the in-cylinder pressure. MPRR is an indicator of engine knocking. As seen in the graph, MPRR values increase with the engine load. As the throttle opens, the volumetric efﬁciency and the total amount of the charge taken into the cylinder increase. Therefore, both in-cylinder pressure and pressure rise rate increase. Octane number of the fuel is one of the most signiﬁcant parameter for MPRR. At the same ignition advance, the fuel having higher octane number provides better antiknock performance. The RON of fusel oil and gasoline used in this study is 106.85 and 96.47 respectively. At full load condition MPRR of gasoline is over 2.5 bar/°CA. However, the octane number of a fuel is not sufﬁcient to explain the MPRR. Flame development and propagation properties of a fuel also have effects on MPRR. MPRR may increase as the burning velocity increases. In this manner, MPRR reﬂects the heat release rate . Flame development and ﬂame propagation durations are seen in Fig. 9 for F0, F50 and F100 fuels at 2500 rpm and four different engine loads. Flame development and ﬂame propagation durations are important parameters inﬂuencing combustion and thermal efﬁciency of the engine. Flame development duration CA0–10 identiﬁes the crank angle duration between spark discharge angle and 10% of mass fraction burnt. CA0–10 duration indicates the mixture ﬂammability which is related to the fuel properties, quality of mixture homogeneous and incylinder temperature before the ignition. Longer CA0–10 durations
Table 3 The properties of test fuels.
Density [kg/m ] Lower heating value [kJ/kg] Motor octane number (MON) Research octane number (RON) Freezing point [°C]
ASTM D 4052 ASTM D 240 ASTM D 2700 ASTM D 2699 ASTM D 6749
746 43,594 86.59 96.47 −52
847 29,536 103.72 106.85 ≤50
785 39,585 89.8 98.7 ≤50
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Fig. 2. Cumulative heat release versus crank angle.
can affect the engine efﬁciency negatively [60,61]. As shown in Fig. 9 CA0–10 durations diminished with the increasing engine load for all types of fuels. Increasing in-cylinder temperatures before the ignition due to engine load enhance the vaporization and homogeneity of the mixture, therefore ﬂame development durations shorten. In general, small portions of oxygenated fuel addition to the gasoline provide shorter ﬂame development and rapid burn than gasoline . The oxygenized hydrocarbon has faster burning rate than gasoline [63–65]. However, CA0–10 duration is prolonged with the usage of the fusel oil at all engine loads. The oxygen availability in the fusel oil may be beneﬁcial to the ﬂame development and ﬂame propagation durations. However, higher latent heat of evaporation and lower LHV of the fusel oil will cause a temperature drop affecting ﬂame development negatively. In addition, the water content of the fusel oil can restrict the ﬂame development. Therefore, CA0–10 durations which can be stated as ﬁrst combustion phase consisting ignition delay are getting longer with using fusel oil. Flame propagation duration CA10–90 identiﬁes the crank angle duration of 10–90% heat release of the total fuel mixture. Flame propagation duration is an important parameter to determine the burning velocity and combustion completeness [66,67]. Longer ﬂame propagation durations cause higher heat losses to the cylinder wall which decreasing engine thermal efﬁciency. The same as CA0–10, CA10–90 durations are also prolonged with the usage of the fusel oil due to the reduced heat release and burning velocity as shown in Fig. 9. As a result, combustion will continue to the end of the expansion stroke. Thereby, effective work and thermal efﬁcieny of the engine will decrease because of the increasing exhaust losses. Extended CA10–90 with using fusel oil is most likely due to the water content and lower LHV of the fusel oil.
Fig. 3. Cylinder pressure and pressure rise rate variation with crank angle.
Fig. 4. Effects of fusel oil on in-cylinder pressure and heat release rate at 2200 rpm engine speed and 25% engine load.
CA50, the crank angle ATDC corresponding 50% mass fraction burnt, is a strong indicator of the engine working conditions. Fig. 10 depicts the variation of CA50 with engine load for F0, F50 and F100 fuels at 2500 rpm engine speed and 20 °CA BTDC ignition advance. In theory, the smallest CA50 provides better combustion and working efﬁciency. Moreover, it is speciﬁed that an ideal engine's CA50 point corresponds TDC. However, in practice, if CA50 is located around 8–10 °CA ATDC, the maximum imep and maximum thermal efﬁciency are obtained [67–69]. As seen in Fig. 10 CA50s closed to the TDC at high loads for all types of fuels. However, CA50s delayed when the fusel oil was used at all engine loads. CA50 of gasoline at full load is 16.2 °CA. CA50s retarded 3.6 and 12.4 °CA when F50 and F100 fuels were used respectively. Combustion losses will increase with retarded CA50. If CA50 shifts toward to the bottom dead center (BDC), after-burning duration will extend during the expansion stroke. Thereby, exhaust losses will increase. It is seen that the fusel usage without optimizing the ignition timing is reduced combustion efﬁciency. It is most likely due to the water content of the fusel oil. However, if an optimization is performed for an ideal ignition timing better combustion characteristics may be obtained for fusel oil. 3.2. Performance and exhaust emissions Fig. 11 shows the engine torque, brake speciﬁc fuel consumption (BSFC) and brake thermal efﬁciency (BTE) variations with the engine loads of 25%, 50%, 75% and 100%. Engine torque increased as the engine
Fig. 5. Effects of fusel oil on in-cylinder pressure and heat release rate at 2200 rpm engine speed and 50% engine load.
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Fig. 6. Effects of fusel oil on in-cylinder pressure and heat release rate at 2200 rpm engine speed and 75% engine load.
Fig. 8. The variations of maximum rate of pressure rise versus engine load.
load was increased. Engine load was controlled by throttle position in gasoline engines. Therefore, opening the throttle causes an increase in the amount of the charge mixture taking into the cylinder and volumetric efﬁciency at the same engine speed. Because of rising energy content of the in-cylinder charge, engine torque increases at high engine loads. Engine torque was slightly decreased with similar trends at all engine loads when fusel oil was used. Engine torque was decreased averagely 2% and 6% respectively when F50 and F100 fuel were used. As seen also from Fig. 11, when the fusel oil was used, BSFC increased considerably. However, as the engine load was increased, BSFC decreased due to the increasing engine power at the same engine speed of 2500 rpm. The differences between BSFC of gasoline and gasoline–fusel oil blends were higher. At low engine load in-cylinder temperatures are comparatively lower than the high loads. In addition to the lower temperatures, the water content and worse combustion characteristics of fusel oil cause a dramatic decrease in in-cylinder temperatures. Because of it, the ﬂame development and ﬂame propagation durations are prolonged and thermal efﬁciency of the engine lows. Therefore, the BSFC of fusel oil is higher than gasoline at lower engine loads. The lower LHV of the fusel oil is another handicap for engine performance. Lower LHV may cause a decrease engine torque and BTE. As a result of this BSFC may increase. As seen in Fig. 11 BTEs were decreased when fusel oil was used. BTE was decreased averagely 3.3% and 8.6% respectively when F50 and F100 fuel were used. As mentioned in the previous section the CA50 has a great effect on BTE. It was thought that the reduction of BTE in case of using fusel oil was resulted from lower LHV of the fusel oil and shifted
locations of CA50s to the BDC. If an optimized ignition timing is determined for fusel oil, better performance results may be achieved. However, the density of the fusel oil is higher than gasoline. This property may provide an advantage in terms of performance. The mass ﬂow rate of the fuel will increase due to the higher density of the fusel oil. Thereby, the negative effect of the lower LHV can be reduced. CO variations with engine loads of 25%, 50%, 75% and 100% for F0, F50 and F100 fuels at 2500 rpm engine speed and λ = 1 are shown in Fig. 12. Generally, CO emissions have a reduction trend when alcohol based fusels are used in internal combustion engines. The reduction of CO results from oxygenated characteristics and well ﬂammability properties of alcohols fuels. In addition, the latent heat of evaporation of alcohol based fuels is higher than gasoline which enables lower intake manifold temperatures and higher volumetric efﬁciency [18,70, 71]. However, in this study, the usage of fusel oil, which is an alcohol based fuel, increased CO emissions. CO emissions averagely increased 6.7% and 21% using F50 and F100 fuels respectively. In-cylinder temperatures during the ﬂame development and ﬂame propagation durations effects oxidation process of the mixture. Lower in-cylinder temperatures aggravate oxidation, thereby especially ﬂame propagation duration extends and combustion cannot complete. In-cylinder temperatures decreased because of the water content of fusel oil. Thus, CO emissions increased rapidly. The same as the CO emissions, HC emissions increased using the fusel oil due to lower in-cylinder temperatures. Fig. 13 shows the HC and NOx emissions variations with engine load at 2500 rpm and λ = 1. An increase in HC emissions up to 10% and 25% was observed with usage of F50 and F100 fuels respectively. Flame propagation ends
Fig. 7. Effects of fusel oil on in-cylinder pressure and heat release rate at 2200 rpm engine speed and 100% engine load.
Fig. 9. Flame development and ﬂame propagation durations of F0, F50 and F100 fuels at 2500 rpm engine speed and four different engine loads.
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Fig. 10. The crank angles for 50% mass fraction burnt of F0, F50 and F100 fuels at 2500 rpm engine speed and four different engine loads.
through the cooler cylinder walls. HC emissions occur at the cooler regions of the cylinder that existing higher heat losses especially cylinder walls [33,72]. NOx emissions decreased because of the decreasing incylinder temperatures when fusel oil was used. NOx emissions decreased 15% and 31% when F50 and F100 fuels were used respectively. NOx formation occurs about 1800 °C in-cylinder temperature. In addition, higher oxygen content in the mixture promotes the NOx formation [4,73]. Water content of the fusel oil reduces in-cylinder temperatures and as a result NOx formation decreases.
Fig. 11. Torque, brake speciﬁc fuel consumption and brake thermal efﬁciency variations of F0, F50 and F100 fuels at 2500 rpm engine speed and four different engine load.
Fig. 12. Effects of fusel oil on CO emissions at four different engine load, 2500 rpm engine speed and λ = 1.
4. Conclusions In this study, effects of the fusel oil on engine performance, emissions and combustion characteristics were investigated. F0 (gasoline), F50 (50% gasoline + 50%fusel oil) and F100 (fusel oil) fuels were tested in a single cylinder, four stroke, port fuel injected research engine. Intake air temperature, ignition timing and λ were kept constant at all experiments. The experiments were conducted at 2500 rpm engine speed and four different engine loads which are 25%, 50%, 75% and 100%. In-cylinder pressures, heat release rates, IMEPs, CA0–10s, CA10– 90s, CA50s, engine torque, BSFC, CO, HC and NOx emissions were examined. Experiments showed that the usage of the fusel oil in the engine aggravated combustion characteristics. Maximum in cylinder pressures, IMEPs and heat release rates decreased when fusel oil was used. Flame development and ﬂame propagation durations were prolonged considerably. CA50s shifted toward to the BDC signiﬁcantly. These results were resulted from the high water content of the fusel oil. In addition, engine performance and exhaust emissions were affected negatively. CO and HC emissions increased 21% and 25% respectively when the F100 fuel was used. On the other hand, as a result of worsen combustion; NOx emissions decreased 31% when F100 was used. As a consequence, it is not possible to use fusel oil directly in a spark ignition engine. The water content of the fusel oil must be removed before using in internal combustion engine. Also,an optimization for ignition timing should be performed for an optimal working. Therefore, combustion and performance characteristics of the fusel oil may be improved.
Fig. 13. Effects of fusel oil on HC and NOx emissions at four different engine load, 2500 rpm engine speed and λ = 1.
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